1. Field of the Invention
The present invention relates to a cylindrical roller bearing. In particular, it relates to a cylindrical roller bearing favorably incorporated in machine tools, engines such as jet engines or gas turbines, and other applications, to support a high-speed rotating shaft.
2. Description of the Related Art
There are many cases where the main spindle assembly of a machine tool such as a machining center, a CNC lathe or milling machine, etc. is driven at high rotation for reasons including improvement of work processing efficiency and precision. There is especially a marked recent trend toward increasing the main spindle rotation speed.
Generally, the main spindle of the main spindle assembly of a machine tool is supported to be freely rotatable with respect to a housing by a rolling contact bearing. The rolling contact bearing is lubricated by oil mist lubrication, air/oil lubrication, jet lubrication, grease lubrication, or other such method in response to the conditions under which it is being used among other factors. Cylindrical roller bearings, angular contact ball bearings, etc. are used for rolling contact bearings.
Cylindrical roller bearings are generally composed of an inner ring having a raceway on an outer circumference thereof, an outer ring having a raceway on an inner circumference thereof, a plurality of cylindrical rollers disposed to roll freely between the respective raceways of the inner ring and the outer ring, and a cage maintaining the cylindrical rollers spaced equally apart in the circumferential direction.
When flange portions are formed at both sides of the inner ring, recess grooves are respectively located at corner portions where flange surfaces of each of the flange portions of the inner ring and the raceway meet. The recess grooves are mainly formed as undercut grooves during grinding of the raceways and the flange surfaces. Chamfers are located at angle portions where roller surfaces and both end surfaces of the cylindrical rollers meet respectively. Further, an axial dimension between the flange surfaces facing in the axial direction is set slightly larger than the length of the cylindrical rollers, and thus a guide clearance is maintained between the cylindrical rollers and the flange portions.
Since the roller surfaces of the cylindrical rollers and the raceways of the races (inner and outer rings) are in line contact, a cylindrical roller bearing such as has been described has a high ability to handle radial load, and is suited for high-speed rotation. In contrast, when compared to a ball bearing, the amount of heat generated during high-speed rotation is large, and particularly, a problem exists where increase in heat generated, wear at portions where there is sliding contact between the cylindrical rollers and the flange portions, and other effects occur easily.
Specifically, the cylindrical rollers have a degree of freedom in inclination as large as the aforementioned guide clearance, and during rotation of the bearing, an unavoidable phenomenon occurs where the axes of the cylindrical rollers incline relative to the axis of the bearing, i.e., skew. When the cylindrical rollers skew, an axial direction component is generated in driving force provided by the raceway on the revolving side, this becoming an axial thrust F to push end portions of the cylindrical rollers against one of the flange portions. Thus, there are instances of frictional resistance increasing at regions where there is sliding contact between the cylindrical rollers and the flange portion, and the end portions of the cylindrical rollers and the flange portion thereby generate heat and suffer wear.
With respect to this problem, various methods of improvement have been suggested in the related art. For example, there are designs where the state of lubrication at the aforementioned portions of sliding contact is improved by designing the recess grooves with a larger height than the height of the chamfers of the cylindrical rollers and forming expanding tapered surfaces on the flange faces with a predetermined angle outward in the axial direction (for example, Japanese Patent Publication No. Sho 58-43609).
Also, there are designs that by having a structure wherein when the cylindrical rollers have skewed, by the outer circumferential edge portions of both end surfaces of the cylindrical rollers contacting with a portion nearer to the base end than the tip edges of the flange surfaces, the edge load of the portions of sliding contact is smaller compared to when the outer circumferential edge portions of both end surfaces of the cylindrical rollers contact with the tip edges of the flange surfaces (for example, Japanese Patent Laid-Open Publication No. Hei 7-12119).
As has been discussed, the cylindrical rollers have a degree of freedom in inclination as large as the guide clearance, and during rotation of the bearing, the cylindrical rollers rotate on their own axes as well as revolve about a shaft while their attitudes are ever-changing within a range of a maximum skew angle θMAX. The maximum skew angle θMAX here means a state where the outer circumferential edge portions of both end surfaces of the cylindrical rollers contact with the flange portions of both sides of the raceway, being at the maximum skew angle within the total degree of freedom of the skew angle of the cylindrical rollers.
As shown in FIG. 10, when cylindrical rollers 23 skew at a skew angle θ less than the maximum skew angle θMAX, the cylindrical rollers 23 are pushed in one direction axially by the previously mentioned axial thrust F, and roll while being guided in a state where they are pushed against one of the flange portions 21b. In this case, the state of contact between the cylindrical rollers 23 and one of the flange portions 21b changes according to the skew angle θ (0<θT<θU<θMAX) in the following manner.
In the range where the skew angle θ is 0<θ≦θT, as shown in FIG. 11, a first boundary R13 between end surfaces 23b and chamfers 23c of the cylindrical rollers 23 contacts a second boundary R11 between flange surfaces 21b1 and recess grooves 21c (contact point shown by a black “●”). Then, in the range where the skew angle θ is θT<θ<θU, as shown in FIG. 12, the first boundary R13 between the end surfaces 23b and the chamfers 23c of the cylindrical rollers 23 contacts the flange surfaces 21b1 (contact point shown by a black “ ”). As the skew angle θ next approaches θU, the first boundary R13 between the end surfaces 23b and the chamfers 23c of the cylindrical rollers 23 contacts with a third boundary R12 between the flange surfaces 21b1 and flange chamfers 21b3 (not shown in the drawings). After this, both end portions of the cylindrical rollers 23 contact with both of the flange portions 21b respectively to attain the maximum skew angle θMAX (not shown in the drawings).
FIG. 13 shows the relation between the skew angle θ of the cylindrical rollers 23 and the pressure P at the contact surface of the cylindrical rollers 23 and the flange portions 21b (solid line), and also shows the relationship of the axial thrust F acting on the cylindrical rollers 23 (broken line). As shown by the same drawing, the axial thrust F becomes larger accompanying an increase in the skew angle θ.
Within the range of 0<θ≦θT, a phenomenon occurs where the pressure P at the contact surface increases at a comparatively steep gradient accompanying an increase in the skew angle θ. This is related to the cylindrical rollers 23 and one of the flange portions 21b contacting at the first boundary R13 and the second boundary R11 (the state shown in FIG. 11), and the axial thrust F becoming larger accompanying an increase in the skew angle θ. Particularly, it has been verified through testing that the pressure P at the contact surface becomes greater than a surface pressure level P0 where wear occurs at the regions of contact thereof in the range θ0≦θ≦θT (region shown by cross-hatching in the same drawing).
When the skew angle θ exceeds θT, the pressure P at the contact surface decreases to a value below the surface pressure level P0 where wear occurs at the portions of contact between the cylindrical rollers 23 and one of the flange portions 21b, after which a stable transition is shown at a comparatively low value irrespective of an increase in the skew angle θ. This is related to a move in the state of contact between the cylindrical rollers 23 and one of the flange portions 21b from contact between the first boundary R13 and the second boundary R11 (the state shown in FIG. 11) to contact between the first boundary R13 and one of the flange surfaces 21b1 (the state shown in FIG. 12).
As the skew angle θ approaches θU, as shown in FIG. 13, the pressure P at the contact surface once again begins a steep increase, and becomes a value exceeding the surface pressure level P0 at the point where θU has been reached. This is related to the state of contact between the cylindrical rollers 23 and one of the flange portions 21b moving from contact between the first boundary R13 and the flange surfaces 21b1 (the state shown in FIG. 12) to a state of contact between the first boundary R13 and the third boundary R12.
As previously explained, the pressure P at the contact surface of the cylindrical rollers and the flange portions is at a value exceeding the surface pressure level P0 where wear occurs at the portions of contact between the cylindrical rollers 23 and the flange portions 21b at the stage before reaching the maximum skew angle θMAX, that is, in the range where the skew angle θ is θ0≦θ≦θT, θU≦θ<θMAX. This is thought to be a major factor contributing to such problems as heat generation and wear at the contact portions.
However, the invention disclosed in Japanese Patent Publication No. Sho 58-43609 previously discussed makes no recognition of the aforementioned phenomenon, and does not suggest a way for improving it. Likewise, the invention disclosed in Japanese Patent Laid-Open Publication No. Hei 7-12119 restricts the state of contact between the outer circumferential edge portions of both end surfaces of the cylindrical rollers and the flange surfaces at the maximum skew angle θMAX, and there is no recognition of the discussed heat generation, wear, or other problems occurring at the stage θ0≦θ≦θT, and does not suggest a means for improvement.
The discussed heat generation, wear, and other problems occurring at the stage θ0≦θ≦θT occur easily in instances where a cylindrical roller bearing formed having flange portions only on one of the inner and outer rings (e.g., NU type, N type, etc.) is driven at high speeds with negative radial internal clearance (preload condition). Cylindrical roller bearings formed having flange portions only on one of the inner and outer rings do not handle axial load, and thus the skew of the cylindrical rollers occurs as a result of such factors as misalignment during bearing installation and slight imperfections in the shape of the raceways of the bearings. The skew angle thereof is therefore only slight with a high probability of being θ0≦θ≦θT, and surface pressure at the contact portions between the cylindrical rollers and one of the flange portions becoming a value larger than the surface pressure level P0 where wear occurs is a frequent state, which has been verified through testing. Further, heat generation, wear, etc. progresses easily when driven at high speeds, since the sliding speed at the portions where there is contact between the cylindrical rollers and the flange portions is high.